Method and Device for Monitoring and Controlling a Hydraulic Actuated Process

ABSTRACT

A method and system for monitoring and controlling a hydraulic pump condition. More specifically, the invention relates to a method and device for monitoring and controlling a hydraulic actuated process indirectly by monitoring and controlling an electric motor driving a positive displacement hydraulic pump. This invention also relates to a precision hydraulic energy delivery system. Direct coupling of the pump to a primary mover (motor) and related motor control allows for complete motion control of a hydraulically driven machine without the use of any downstream devices. By employing motion control algorithms in the motor control, the hydraulic output at the pump head is controlled in a feed forward method.

INTRODUCTION

1. Field of the Invention

The present invention relates to a method for monitoring and controllinga hydraulic pump. More specifically, the invention relates to a methodand device for monitoring and controlling a hydraulic actuated processindirectly by monitoring and controlling an electric motor driving apositive displacement hydraulic pump.

This invention also relates to a precision hydraulic energy deliverysystem. Direct coupling of the pump to a primary mover (motor) andrelated motor control allows for complete motion control of ahydraulically driven machine without the use of any downstream devices.By employing motion control algorithms in the motor control, thehydraulic output at the pump head is controlled in a feed forwardmethod.

2. Background of the Invention

In the prior art, it is well known that in situations where higherpressures of fluid movement are desired, a positive displacement pump iscommonly used. A positive displacement pump is usually a variation of areciprocating piston and a cylinder, of which the flow is controlled bysome sort of valving. Reciprocal machinery, however can be lessattractive to use than rotary machinery because the output of areciprocal machine is cyclic, where the cylinder alternatively pumps orfills, therefore there are breaks in the output. This disadvantage canbe overcome to a certain extent by: using multiple cylinders; bypassingthe pump output through flow accumulators, attenuators, dampers; orwaste gating the excess pressure thereby removing the high pressureoutput of the flow.

In addition to uneven pressure and flow output, reciprocating pumps havethe disadvantage of uneven power input proportional to their output.This causes excessive wear and tear on the apparatus, and is inefficientbecause the pump drive must be sized for the high torque required whenthe position of the pump connecting rod or cam, in the case of an axial(wobble plate) pump, is at an angular displacement versus the crankarmdimension during the compression stroke that would result in the highestrequired input shaft torque.

Moreover, if the demand of the application varies, complicated bypass,recirculation, or waste gate systems must be used to keep the systemfrom “dead-heading.” That is, if flow output is blocked when the pump isin operation, the pump will either breakdown by the increased pressureor stall. If stalling occurs, a conventional induction electric motorwill burn out as it assimilates a locked rotor condition with full ratedvoltage and amperage applied. Typically systems with fixed displacementpumps use a relief valve to control the maximum system pressure whenunder load. Therefore, the pump delivers full flow at full pressureregardless of the application thus wasting a large amount of power.

In this regard, certain prior art that attempts to correct the problemsassociated with torque output of a pump motor should be noted.

In U.S. Pat. No. 5,971,721, an eccentric transmission transmits a torquedemand from a reciprocating pump, which varies with time, to the drivemotor such that the torque demand on the drive motor is substantiallyconstant. The result is the leveling of torque variation required todrive a positive displacement pump at the transmission input shaft withthe effect of constant pump output pressure. This is accomplished bymeans of eccentric pitch circle sprocket sets with gear belts oreccentric pitch circle matched gear sets.

The use of an eccentric gear or sprocket set has a significant effect onthe overall torque requirement and the magnitude of the discharge pulseof the pump. But, because most pumps are of a multi-cylinder or are vaneor gear types, the pump input shaft torque requirement would not beperfectly counter-acted (leveled) by using the reduction patterndeveloped by eccentrically matched transmission components.

In U.S. Pat. No. 5,947,693, a position sensor outputs a signal bysensing the position of a piston in a linear compressor. A controllerreceives the position signal and sends a control signal to controldirectional motion output from a linear motor.

In U.S. Pat. No. 4,726,738, eighteen or nineteen torque leads aremeasured along the main shaft in order to maintain constant shaftvelocity revolution and are translated to a required motor torque forparticular angles of the main shaft.

U.S. Pat. No. 4,971,522 uses a cyclic lead transducer input andtachometer signal input to a controller to signal varied cyclic motorinput controls to provide the required motor torque output. A flywheelis coupled to the motor in order to maintain shaft velocity. However,the speed of the motor is widely varied and the torque is varied to asmaller extent.

U.S. Pat. No. 5,141,402 discloses an electrical current and frequencyapplied to the motor which are varied according to fluid pressure andflow signals from the pump.

U.S. Pat. No. 5,295,737 discloses a motor output which is varied by acurrent regulator according to a predetermined cyclic pressure outputrequirement. The motor speed is set to be proportional to the volumeconsumed and inversely proportional to the pressure.

It is seen from the foregoing that there is a need for electronicattenuation of the torque profile in a pump. When the torque profile iscompared with the input shaft displacement and other known factors suchas system inertia and response time of the pump drive etc., a pump canproduce constant pressure and therefore constant flow without thetypically associated ripple common to power pumps for the full range ofthe designed volumetric delivery, by driving them in a feed forwardmethod.

It should be noted that the foregoing hydraulic pumping systems controloutput pressure and flow in the micro sense. These concepts examinemodulating the input shaft torque and speed to provide a constanthydraulic output, whether it is pressure or flow limited. See U.S. Pat.No. 5,971,721, U.S. Pat. No. 6,494,685, and U.S. Pat. No. 6,652,239, theentire contents of which are hereby incorporated by reference.

It should be further noted that attempts to provide a high dynamic rangeof hydraulic flow and pressure during operation of prior pumping systemsrequired placement of downstream devices in the liquid path to modulatethe hydraulic output. With such systems, the pump provides the maximumhydraulic flow (as the prime mover) and the downstream devices adjustthe output to match the application requirements.

The prime mover in such systems is typically a constant speed inductionmotor. In to order to control the hydraulic output, feedback devices, aprocessor (be it mechanically balanced or electronic) and hydraulicservo valves must be placed into the hydraulic stream for flow andpressure regulation. This treatment of hydraulic delivery places the“smarts” of the system in the hydraulic output portion of the system.Disadvantageously, these systems require many hydraulically drivendevices, are mechanically (geometry) limited, are energy inefficientwhen total system performance is scrutinized and have a small range ofdynamic response (typically 10-1). Typical examples of acommand-response curve of a small servo valve and a large servo valveare shown in FIGS. 9 and 10, respectively.

Moving the “smarts” directly into the prime mover—by incorporatingvariable speed (VFC) controlled motors—has been attempted. However, thisprovides limited torque delivery potential at low speeds, and manyfeedback devices are required for its operation. Further, the responseof such a system is only generally higher than the 150 ms range and theenergy savings potential is only in the 50% range.

These approaches address—in the macro sense—the need for a prime movercoupled to a power pump that controls the energy, and therefore the flow(velocity) and pressure (torque) at the input shaft of the pump.Moreover, the desired system must replicate the motion controlcapabilities of existing systems without requiring the use of downstreamflow control devices and feedback circuits.

An example of another hydraulic process, also to which one embodiment ofthe present invention is directed to, would be presses (such a metalforging presses and laminating presses) driven by hydraulic cylinders.In critical applications, valves and feedback controls are used tocontrol the critical process variables of applied force and the rate ofactuation. A variety of transducers are employed both to facilitatecontrol and to monitor the process.

Additionally, level switches, pressure transducers and flow switches areoften employed to monitor system conditions such as low oil levels,plugged pump inlet filters and burst hoses. Use of such transducers addsto the system cost and complexity and is prone to failure particularlyin hostile environments.

Yet another example of a prior art implementation are computercontrolled servo valves relying on feedback from pressure transducers,linear position transducers, and flow meters. The limitations oftraditional systems include reliability, slow response time and controlprecision. A recent research paper in the Journal of Materials titled:Modeling and Simulating Metal-Forming Equipment, by Frazier, Medina,Mullins and Irwin describes the current art as it is applied to forgingpresses. The paper described a metal forging process, which included acylindrical upsetting of plain carbon steel. The press was programmed toforge at a constant velocity of 1.27 cm/sec. A plot of an experimentaland simulated ram velocity profile is shown in FIG. 12. The data clearlyreveals the ramping up and overshoot of the desired velocity. A briefchange in velocity due to impact with the workpiece is observed near 3.5seconds.

FIG. 13 shows an experimental and simulated result for the ram load asderived from the ram pressure measurement. The plot reveals thatapproximately 88.96 kN are needed to overcome the counter-balance andfrictional forces. The load increases rapidly beginning at approximately3.5 seconds. This corresponds to impact with the workpiece. Beginning atapproximately 4 seconds, elastic deformation of the workpiece andtooling ceases and plastic deformation of the workpiece begins. It isclear from the data that the press was able to maintain the desiredvelocity under load, as long as the load did not increase too rapidly,but as soon there was a sudden increase in load, the desired velocity islost.

It is to be noted that a particular valve command does not always resultin the same flow rate. This is due to the changing pressure across theservo manifold, clearly revealing that the flow rate can decrease evenas the command for higher velocity increases.

As the need for precise control of ram velocity, as well as position,increases, the need for direct measurement of velocity as a feedbackcontrol signal will become more acute. Therefore, current techniques forcontrolling velocity (inversion of servovalve flow models combined withpressure measurements and numerical differentiation of displacementmeasurements) are not adequate for high performance over a broad rangeof forging conditions.

Experience shows the need for press operators to customize the presscontrol law in order to achieve the desired velocity profile fordifferent forming operations. Having a custom control law for equipmentthat repeatedly makes the same part for several weeks or more issatisfactory, but as the need to use the same equipment for severaldifferent parts in a day or custom small lots increases, having torepeatedly tune the control law can waste significant time. Controllaws, therefore, need to be designed to be robust so that differentloading conditions and velocity profiles can be handled successfullywithout the need for customized tuning.

The paper's conclusion includes the following statement:

-   -   Current techniques for controlling velocity (inversion of        servovalve flow models combined with pressure measurements and        numerical differentiation of displacement measurements) are not        adequate for high performance over a broad range of forging        conditions.

Accordingly, there is a need for a system and method of eliminating theneed for servovalves and transducers while providing more precisecontrol of both hydraulic press ram force and displacement. There isalso a need for a pump and motor assembly with constant pressure outputand a motor controller with electrical regeneration. With thepossibility of directly sensing workpiece conditions during formingoperations, it is demonstrated that these measurements could be fed backto the metal-forming equipment for computer control of the formingequipment, thereby enabling real-time compensation for variations ininitial workpiece and equipment conditions. From a process controlperspective, this approach enables the highest level of robustness andrepeatability in production. Common to each of these factors is the needfor improved control and predictability of the equipment's behavior,which itself is based on the desire to achieve near-net shapes,higher-quality end products, higher yields, and better control ofmicrostructure, especially for hard-to-form materials in metal forgingprocess.

More specifically, the current synchronized motion control methods usedin hydraulic applications are not designed to possess the dynamichydraulic properties which affect each axis independently that areresponsible for coordinating force inherent in classical multi-axis,hydraulic machinery, such as for example shown in FIG. 15. Consequently,these controllers cannot easily maintain coordination for all operatingconditions without expensive and complicated feedback devices andcontrol loops.

SUMMARY OF THE INVENTION

It is therefore an object of the invention to device a control methodfor use in hydraulic applications that possesses the dynamic hydraulicproperties affecting each axis independently that are responsible forcoordinating force inherent in classical multi-axis and multi-primemover, hydraulic machinery.

A further objective of the invention is to provide controllers thatmaintain coordination for all operating conditions without expensive andcomplicated feedback devices and control loops.

Another object of the present invention is to provide a method forelectronic attenuation of pump torque variation requirements in order toproduce a matched motor torque output that will result in constantoutput pressure from a pump.

Yet another object of the present invention is to provide monitoring andcontrol factors which vary the power and torque output of a pump motorbased on calculated torque variation requirements.

Yet another object of the present invention is to increase the energyefficiency of a pump system, by providing a force balanced relationshipbetween the motor output and the application's hydraulic requirement,thus allowing the use of energy saving torque drives without incurringthe pressure variations associated with their use.

Yet another object of the present invention is to decrease the wear andtear on the pump by providing a substantially constant force output fromthe motor of the pump and reduce the amount of cycles of the pump to theapplication's requirement.

Yet another object of the present invention is to provide a method forelectronic attenuation of pump torque variation by supplying informationfor design of an electronic transmission system that can achieve amodulated torque output from the motor to the pump.

Yet another object of the present invention is to achieve precisecontrol of flows and pressures, thus precise control of press ramvelocities, force and position.

Yet another object of the invention is to create precise hydraulicoutput from the pump utilizing an algorithm programmed into the drivecontrol that systematically measures and corrects for three key physicalparameters of the motor, pump and hydraulic fluid combination. The threeparameters being windage torque, viscous torque and coulomb torque. Oneadvantage provided by this aspect of the invention is that the resultantprecision allows multiple (two or more) drive, motor and pumpcombinations to be electronically “line-shafted” together whileproviding precise and stable control of pressure and flow being fed to acommon output header regardless of individual pump characteristics orfluid condition variations.

The invention also provides for development of an empiricalunderstanding of the positive displacement pump in regards to slippage(variation from theoretical displacement) throughout the fullpressure/flow delivery range of the pump. This understanding combinedwith the drive controller's precise measurement of motor rpm gives anextremely precise measurement and control of hydraulic fluid flow rate.

In turn, knowing the displacement volumes of the cylinder being actuatedand mathematically correcting for elasticity of the piping system allowsfor very precise control and or measurement of press-ram position,and/or ram velocity than may be achieved with prior art.

The invention employs the feed forward torque control aspects of thedrive system to precisely monitor and/or control hydraulic pressurescreated by the pump which after mathematical correction for line lossesat various flow rates, mimics load at the press ram. The invention alsoprovides both a more precise and a more robust method than does theprior art.

To attain the objects described, there is provided a method forobtaining a polar map for process control within the electronic drive ofa targeted pump. This polar map is calculated by a processor or isexternally calculated then input into a processor. Once the torqueprofile of the pump is obtained and translated into a polar map, theprocessor can compare the shaft displacement angle of the pump inputshaft to the reference polar map. The processor can also take intoaccount selected factors such as the response time of the pump drive,the motor inductive reactance, system inertia, applicationcharacteristics of the pump, and regenerative energy during decelerationof the pump.

Using selected factors and the comparison results, the processor signalsthe motor controller to vary the amperage, voltage, and frequencyapplied to the motor in order to regulate the torque output of the pumpmotor. With an accurately modulated motor torque output in concert withthe established polar map (for the targeted pump), the pump outputpressure will remain constant regardless of the pump's crank armlocation or the velocity of fluid flow.

It is also an object of the present invention to provide a hydraulicenergy delivery system that allows for complete motion control of ahydraulically driven machine with the use of minimal or no downstreamfeedback devices.

It is therefore a further object of the present invention to providecontrol factors which vary the power and torque output of a pump motorby employing motion control algorithms.

To attain the objects described, there is provided direct coupling of apositive displacement pump to a pump drive motor and related controls.By employing motion control algorithms into the motor control, thehydraulic output at the pump head will simultaneously follow. Controlfeatures listed herein may be integrated into the system by developingalgorithms and subroutines for the control system coupled to the pump.

These and other objects and advantages are provided by the presentinvention, preferred embodiments of which are described in theparagraphs below.

All features listed herein may be integrated into the system bydeveloping algorithms and subroutines for monitoring and controlling thesystem coupled to the hydraulic pump.

The present invention will now be described in more complete detail withreference being made to the figures identified below.

BRIEF DESCRIPTION OF THE DRAWINGS

The following detailed description, given by way of example and notintended to limit the present invention solely thereto, will best beappreciated in conjunction with the accompanying drawings, wherein likereference numerals denote like elements and parts, in which:

FIG. 1 is a block diagram of the steps required for a method ofelectronic attenuation of torque profile and the resulting control ofthe pump;

FIG. 2 is a graph depicting individual input torque variation for eachnode of a triplex pump based upon pump input shaft rotational degrees;

FIG. 3 is a graph depicting a percentile summation of input torquevariation compared to angular displacement of the input shaft of atriplex pump;

FIG. 4 is a table depicting variations of input torque above and belowthe mean for triplex pumps in relation to the linear distance betweenthe plunger/piston pivot point and the throw pivot point multiplied bythe throw radius;

FIG. 5 is a graph depicting a plotting of geometric distance variationpoints based upon the summation of total torque variation for a triplexpump;

FIG. 6 is a polar map depicting the torque profile versus angulardisplacement of a pump input shaft;

FIG. 7 is a diagram illustrating a precision hydraulic delivery systemaccording to the present invention;

FIG. 8 is a graph depicting a profile of torque vs. velocity for anexemplary hydraulic system in accordance with the present invention;

FIG. 9 is a graph depicting command, response curves of a small servovalve;

FIG. 10 is a graph depicting command, response curves of a large servovalve;

FIG. 11 is a schematic of a hydraulic press system according to oneembodiment of the invention;

FIG. 12 is a graph depicting commanded, simulated and experimental ramvelocity profiles not within normal boundaries;

FIG. 13 is a graph depicting simulated and experimental ram loadprofiles not within normal boundaries;

FIG. 14 is a diagram of “electronic line-shafting” with a master/slaveconfiguration of multi-drop positive displacement hydraulic pumps,according to one embodiment of the invention;

FIG. 15 is a schematic of a typical non-compensated electronicline-shafting in a multi-axis hydraulic machinery; and

FIG. 16 is a diagram of a system incorporating the Learn TQ items:Coulomb torque, Windage, viscous coefficient and applicationcharacteristics: Redux P, Redux V feed-forward compensation algorithm,according to one embodiment of the invention.

The description of the various elements of the invention will bediscussed in detail in the following sections.

DETAILED DESCRIPTION OF THE INVENTION

The instant invention will now be described more fully hereinafter withreference to the accompanying drawings, in which preferred embodimentsof the invention are shown. This invention may, however, be embodied inmany different forms and should not be construed as limited to theillustrated embodiments set forth herein. Rather, these illustratedembodiments are provided so that this disclosure will be thorough andcomplete, and will fully convey the scope of the invention to thoseskilled in the art.

Referring now to the drawings in detail wherein like numerals refer tolike elements throughout the several views where Blocks 1-5 of FIG. 1depict the development of a baseline polar guide of the torque profilefor the targeted pump.

In Block 1 of FIG. 1 and graphically depicted in FIG. 2, the outputcharacteristic of volumetric displacement would directly relate to theinput torque variations above 10 and below 12 the comparative mean 14.The processor identifies the output discharge characteristics such asthe number of plungers, pistons in a piston pump, or vane/gear in arotary pump. The processor also utilizes a comparative mean where, thecomparative mean is representative of the basic torque requirement ofthe pump input shaft rated at a specific output pressure of the pump. Apulsation pattern 16 would be repeated at the same rate per revolutionas the number of the pump's volumetric displacement cavities. Asillustrated in FIG. 2, a triplex positive displacement pump would repeata pulsation pattern 16 every 120 degree rotation of the pump inputshaft. These torque variations above 10 and below 12 the mean 14 arecalculated and recorded for Block 1 of FIG. 1.

For other pumps such as a quintaplex plunger pump, which incorporatesfive plungers, a pulsation pattern would be produced five times perrevolution of the pump input shaft, repeating every 72 degrees if theoutput pressure is to remain constant; and for a rotary vane pump withnine vanes selected, the pulsation pattern would repeat every 40 degreerotation of the pump input shaft if the output pressure is to remainconstant.

In Block 2 of FIG. 1 and depicted graphically in FIG. 3, the torqueprofile versus displacement angle of the targeted pumping system is thesummation of the torque requirement for each volumetric displacementcomponent, depicting a percentage above mean 18 and the percentage belowmean 20.

In Block 3 of FIG. 1, the magnitude of the input torque variation forthe power pump is determined by the processor, where the magnitude ofthe torque variation is the number of volumetric displacement cavitiesactivated in one revolution and the relationship “Q”. The calculation“Q” is the linear distance “L” between the plunger/piston pivot pointand the throw pivot point multiplied by the throw radius “R”; “Q=LR”.FIG. 4 in table form, depicts the percentile variations of input torqueabove and below the mean for triplex pumps with various “Q”.

FIG. 5 graphically depicts the total torque variation to show a torqueprofile for a triplex pump (three volumetric displacements perrevolution) with a “Q” at 4:1 with variations shown above and below themean. The mean is representative of the basic rms (root mean squared)torque requirement of the pump input shaft rated at a specific outputpressure of the pump versus the angular displacement of the pump crankshaft. The relationship of “Q” and the effect it has on torque variationwould also apply to rotary pumps. A plotted geometric distance variationusing t1-t15 (as plotting points) is then imposed on the torque profile.

In Block 4 of FIG. 1 and graphically depicted in FIG. 6, a pump polarmap is determined based on the torque profile and the input shaftangular displacement of the pump. The center 34 of the polar map is torepresent zero torque. The incremental lines 36 depicted orbitally arethe angular displacement of the targeted pump's input shaft. The plottedpump torque variation curve 38 that occurs above and below the mean 40is to be considered a geometric percentage of the summation of thetorque requirement of each of the volumetric displacement components ofthe targeted pump.

The distance of each point plotted on the polar map's center from thebase diameter's center is the geometric distance variation (over orunder) of the base radii percentile established from torque versus thepump input shaft displacement angle (t1 thru t15). The geometricdistance variations are the plotting points determined in FIG. 5. Thetorque versus angular displacement profile of the pump system selectedis to become the reference polar guide for the comparitor algorithm inthe processor in Block 5 of FIG. 1. The reference polar guide determinedby the processor in Blocks 1-5 can also be determined externally fromthe processor and then input into the processor.

Blocks 6-10 of FIG. 1 are the operating steps from electronicattenuation of the torque profile to provide a constant output pressureat the pump, wherein Block 6 indicates the transmission of the angulardisplacement of the input shaft of a pump in operation. A pulsetransmitter mounted on the input shaft relays to a counter—which is partof the processor—the angular position of the pump drive.

In Block 7 of FIG. 1, an electronic processor gathers this output shaftorientation feedback information, and processes the angular displacementdata. The processor then attenuates from the peak requirement of thepump, the output torque of the drive compared to the predeterminedreference polar map of Block 5. A corresponding torque command value isthen selected.

In Block 8 of FIG. 1, other inputs of system readings such as systeminertia, parasitic leads, off throttle friction, response time of thepump, motor inductive reactance, application characteristics of thepump, regenerative energy during deceleration of the pump, andtranslation speed can be selectively factored into the processoralgorithm for changes in process control.

In Block 9 of FIG. 1, based upon the inputs of Blocks 7 and 8, theprocessor of the electronic drive signals the motor controller to applythe correct amperage, voltage, and frequency to the motor which thenprovides the correct torque according to the angular displacement of thepump input shaft.

In Block 10 of FIG. 1, the resultant signal to the motor controller andmotor will drive the pumping system to produce constant pressure at thefull range of the designed system flow volume regardless of pump radialcrankshaft location and the velocity of the fluid pumped.

Block 11 of FIG. 1, depicts the use of this method in future systemswhere information gathered from pump operation by this method can beused to design more responsive components such as transmissions andelectronic drives. More responsive components would decrease the timeincrements between Blocks 6-10. As response times are decreased, thetorque output produced for indicated angular displacements will increasein efficiency.

FIG. 7 depicts a precision hydraulic delivery system 71 according to thepresent invention. Advantageously, this system provides direct couplingof a positive displacement pump 72 to a prime mover 73 and related motordrive control 74. The prime mover 73 in the pump system shown is, forexample, a constant speed induction motor. The motor has, for example, a1000-1 (torque) turn down ratio. The motor control 74 may be, forexample, an electronic servo or hydraulic type motor control. Directcoupling of the pump 72 to the motor 73 and motor control 74 allows forcomplete motion control of the pump 72 without requiring any of thedownstream flow control devices, feedback devices, hydraulic energystorage devices (accumulators) or energy dissipation devices normallyused in conventional pump systems.

The system in FIG. 7 employs motion control algorithms in the electronicmotor control so that the hydraulic output at the pump head willsimultaneously follow the control signals generated by the algorithmsand sent to the motor. This ability allows a large dynamic range ofhydraulic energy to be delivered by placing the “smarts” of the systemdirectly into the electrical handling capabilities of the prime movercircuit. The modulation of torque (resulting in hydraulic pressure) andvelocity (resulting in hydraulic flow) are most efficiently handledwithin the electronic servo or hydraulic type control of the primarymover.

The teachings of U.S. patent application Ser. No. 09/821,603 and U.S.Pat. No. 5,971,721, which are hereby incorporated by reference, may beincorporated into the macro motion control capabilities described hereinto provide improved system response, “keypad” tuning of a hydraulicapplication, very high systemic efficiency characteristics andsimplified hydraulic circuitry.

Several exemplary control features of the present invention aredescribed in greater detail below. These features represent only afraction of the possible features that may be electronically integratedinto a hydraulic delivery system by control algorithms and subroutinesfor a prime mover servo control system coupled to a pump.

“SLAM Absorption” Feature

The “SLAM” subroutine is an energy absorbing function that provideshydraulic component protection by eliminating pressure spikes. In someapplications, a “spike” in pressure occurs when flow volume is rapidlyreduced. This normally occurs when, for example, a directional controlvalve is shut, and is typically followed by the pressure relief valvewaste-gating the excess flow to a tank until the system flow returns tonormal.

This condition is undesirable, and to eliminate it the present inventionhas a discrete input that activates the “SLAM” function when such anevent occurs. A determination as to the likelihood of such an event ismade during commissioning. Use of the “Position Sensing” feature(described below) allows the “SLAM” subroutine to be invoked whennecessary. The “SLAM” feature causes the electronic drive to capture theinertial energy of the system via the regenerating capabilities of theprime mover (turning the motor into a generator), and to store thiscaptured electrical energy in the electronic drive (see “energy storagesystem” below). The normally waste-gated energy is thus captured by thedrive during this function, thereby saving energy and reducing wear onthe hoses and hydraulic system.

“JAB Applied” Feature

The “JAB” feature eliminates pressure “droop” by invoking a rapid pumpacceleration feature of user defined time and amplitude, that is appliedover and above the normal flow or pressure input commands. In someinstances, a rapid increase in flow volume required by the applicationwill cause the pressure to droop until high inertia components in thepumping system are accelerated to the required delivery velocity. Ifthis droop is undesirable in a specific application, a discrete inputcan be used to activate this “JAB” rapid acceleration feature that isapplied over and above the normal flow or pressure input commands thatare controlling the pump.

Dual Function Pump/Motor Feature

This feature provides for single unit hydraulic motor/pump functionsfrom the same hydraulic device for energy delivery and reclamation(regeneration and storage).

“Pressure Loop” Feature

This feature provides a pump shaft torque output measurement methodwhich is translated into a pressure delivered signal.

“Constant HP System” Feature

This feature provides a constant horse power electrical drive system formaintaining an energy ceiling regardless of the delivered flow volume.

“Energy Storage System” Feature

This feature provides an electrical energy storage device in the drivesystem for reclamation of energy from regeneration (see “Dual functionpump/motor” and “SLAM” function), or for high output energy spikestypically provided by a hydraulic accumulator. “Position Sensing”feature

According to this feature, a volumetric pulse correlates to a pumpoutput volume that will cause an incremental pulse to occur. Thisvolumetric pulse (output by the electronic drive module inclusive ofcompensation factors Pump TQ Learn and application items REDUX V andREDUX P) is used for the positioning of known hydraulic cylinders andtheir corresponding volumetric displacements.

“Leakage Detection” Feature

This subroutine is used to detect user defined excessive hydraulicleakage rates. This feature compares the output of the “PositionSensing” function to a known limit during a move, and if there is adiscrepancy beyond a predetermined amount, an alarm output results.

“Output Gain Offset” Feature

This feature allows the user to assess the output gain levels of thehydraulic delivery (pressure vs. flow) in order to overcome anyapplication flow restrictions or mechanical variation. The assessmentresults in a profile of torque vs. velocity for the desired hydraulicoutput.

FIG. 8 shows an example 5 point torque profile, including:(1) Gain Zero801, (2) Gain Lo 802, (3) Gain Mid 803, (4) Gain Hi 804, and (5) GainMax 805. The five gain points plotted on the graph are described below.

1. Gain Zero: For “pressure delivered” vs. “zero velocity” (the RPM ofthis point is always anchored at zero RPM), the Gain Zero corrects thepressure reference command as the velocity decreases to “0” tocompensate for systemic “sticktion”.

2. Gain Low: For “pressure delivered” vs. “velocity,” the Gain Lowcorrects the pressure reference command as the velocityincreases/decreases to compensate for system losses. Gain Low RPM:Applies the “GAIN LOW” value when the pump system is operating within auser defined RPM range (typically, 0 to 50 RPM). The gain is applied asa tapered offset beginning with the “GAIN ZERO” value at 0 RPM, andending with the “GAIN LOW” value at the “GAIN LOW RPM.” Any operationabove this speed is ramped to the “GAIN MID” point.

3. Gain Mid: For “pressure delivered” vs. “velocity,” the Gain Midcorrects the pressure reference command as the velocityincreases/decreases to compensate for system losses.

Gain Mid RPM: Applies the “GAIN MID” value when the pump system isoperating within a user defined RPM range (typically, 50 to 700 RPM).The gain is applied as a continued offset beginning with the “GAIN LO”value at the “GAIN LO RPM” and ending with the “GAIN MID” value at the“GAIN MID RPM.” Any operation above this speed is ramped to the “GAINHI” point.

4. Gain High: For “pressure delivered” vs. “velocity,” the Gain Highcorrects the pressure reference command as the velocityincreases/decreases to compensate for system losses.

Gain High RPM: Applies the “GAIN HIGH” value when the pump system isoperating within a user defined RPM range (typically, 701 to the maximumRPM). The gain is applied as a continued offset beginning with the “GAINMID” value at the “GAIN MID RPM” and ending with the “GAIN HIGH” valueat the “GAIN HIGH RPM.” Any operation above this speed is ramped to theGAIN MAX RPM point.

5. Gain Max: For pressure delivered vs. DRIVE SPEED MAX velocity (theRPM of this point is always anchored at the drive speed max RPM), theGain Max attenuates the pressure reference command as the velocityincreases/decreases to compensate for system losses.

The invention according to one embodiment of the invention is a methodfor monitoring and controlling a hydraulic pump driving a metal forgingpress. A simplified block diagram of a hydraulic press system 110, forexample a metal forging press, is shown in FIG. 11. The system 110 ispowered by an electric motor that drives a hydraulic pump 118. Transientdemands for high ram speeds are met by an accumulator system 120. Acounter-balance is employed to support and return the main ram 112 tothe top of its stroke after the completion of a forging operation. Thehydraulic manifold 114 controls the flow of fluid to the main ramcylinder 116 and to the tank 122. The metal forging can be a cylindricalupsetting of, for example, plain carbon steel. The press can beprogrammed to forge at any constant velocity, for example at 1.27cm/sec.

One embodiment of the invention is an application of an “electronicline-shafting” control technique which serves to replicate and evenimprove the historical, hydraulic multi-axis coordinated motion controltechniques. This technique incorporates a method of servo-drivenhydraulic prime-mover control on multi-axis hydraulic applications withthe Learn TQ, Redux V, and Redux P compensation factors incorporated fordirect feed forward precise hydraulic output without the need forexternal feedback devices. Redux P relates to the compensation orreduction or compression of a fluid or system capacitance, while Redux Vrelates to the restrictive flow of the pump. The result demonstratesthat the “electronic line-shafting” technique significantly improves thecoordination, robustness, and overall stability of hydraulic poweroutput subjected to realistic physical limitations.

Therefore, key aspects of the present invention include precise controlof press ram velocities, force and position requires precise control offlows and pressures. To create precise hydraulic output from the pump,the invention utilizes an algorithm programmed into the drive controlnamed Pump Torque Learn (abbreviated Learn TQ) that systematicallymeasures and corrects for three key physical parameters of the motor,pump and hydraulic fluid combination. These three factors are accountedfor to understand the relationship between applied torque and resultantpump pressure; a relationship that changes over the range of operatingspeeds of the pump. These parameters are windage torque, viscous torqueand coulomb torque, respectively. A separate advantage provided by thisaspect of the invention is that the resultant precision allows multiple(two or more) drive, motor and pump combinations 125 to beelectronically “line-shafted” together while providing precise andstable control of pressure and flow being fed to a common output header130 regardless of individual pump characteristics or fluid conditionvariations. A multi-drop set-up for positive displacement hydraulicpumps used in “line-shafting” hydraulic supplies, according to oneembodiment of the invention is shown in FIG. 14. A further example of asystem 150 incorporating the Learn TQ, Redux P, Redux V feed forwardcompensation algorithm according to one embodiment of the invention isshown in FIG. 16.

The invention also incorporates development of an empiricalunderstanding of the positive displacement pump in regards to slippage(variation from theoretical displacement) throughout the fullpressure/flow delivery range of the pump. This understanding combinedwith the drive controller's precise measurement of motor rpm gives anextremely precise measurement and control of hydraulic fluid flow rate.In turn, knowing the displacement volumes of the cylinder being actuatedand mathematically correcting for elasticity of the piping system allowsfor very precise control and or measurement of press-ram position,and/or ram velocity than may be achieved with prior art.

As explained above, one embodiment of the invention employs the feedforward torque control aspects of the drive system to precisely monitorand/or control hydraulic pressures created by the pump which aftermathematical correction for line losses at various flow rates, mimicsload at the press ram. The invention therefore provides both a moreprecise and a more robust method than does prior art.

Accordingly, some of the benefits derived from the present inventionare:

1. More precise control of ram velocity profile than prior art.

2. Without relying on external sensors mounted at the process but ratherusing motor current signature analysis combined with motor shaft encoderdata, the drive controller may be taught to recognize, report, and/ortake action against process anomalies. Such anomalies would showthemselves as velocity (flow rates), current draws and torques(pressures) that are deemed to be outside normal process parameters.Examples include:

a) A plugged inlet filter would cause a low current signature combinedwith a high velocity (motor rpm) than would be expected for the flow andpressure being delivered.

b) A burst hydraulic hose would result in an abrupt shift to anabnormally high flow rate.

c) Metal being forged that was not within proper metallurgyspecifications or a billet not heated to the desired temperature wouldproduce ram speed, displacement and force characteristics that thecomputer could recognize or simply log as being not within normalboundaries (such as shown in FIGS. 12 and 13).

The ability to “electronically line-shaft” two or more pump, motor, anddrive systems to feed a common hydraulic output header while maintainingprecise and stable control of both total flow and common pressure.

Thus, while fundamental novel features of the invention are shown anddescribed and pointed out, it will be understood that various omissionsand substitutions and changes in the form and details of the devicesillustrated, and in their operation, may be made by those skilled in theart without departing from the spirit of the invention. For example, itis expressly intended that all combinations of those elements and/ormethod steps which perform substantially the same function insubstantially the same way to achieve the same results are within thescope of the invention. Moreover, it should be recognized thatstructures and/or elements and/or method steps shown and/or described inconnection with any disclosed form or embodiment of the invention may beincorporated in another form or embodiment. It is the intention,therefore, to be limited only as indicated by the scope of the claimsappended hereto.

1. A method for monitoring and controlling a hydraulic actuated process,comprising the steps of: monitoring and controlling an electric motordriving a positive displacement hydraulic pump driving said hydraulicactuated process, said monitoring and controlling being carried out byutilizing an algorithm programmed into the electric motor drive controlthat systematically measures and corrects for any combination of or allthree physical parameters of the motor, pump and hydraulic fluid,thereby optimizing an applied torque and resultant pump pressure, andwherein said three physical parameters are windage torque, viscoustorque and coulomb torque, respectively.
 2. The method according toclaim 1, wherein a plurality of drives, motors and pumps areelectronically line-shafted together providing precise and stablecontrol of pressure and flow being fed to a common output headerregardless of individual pump characteristics or fluid conditionvariations.
 3. The method according to claim 1, wherein a precisemeasurement and control of press-ram position, and/or ram velocity isachieved.
 4. The method according to claim 1, wherein said methodmonitors and/or controls hydraulic pressures created by the pump, whichafter mathematical correction for line losses at various flow rates,mimics load at the press ram.
 5. The method of claim 1, wherein thealgorithm includes a subroutine for eliminating a pressure spike;wherein any excess hydraulic output is used to generate electricalenergy when a pressure spike occurs; the electrical energy being storedin an energy storage means.
 6. The method of claim 1, wherein thealgorithm includes a subroutine for eliminating a pump pressure droop;wherein an input signal overrides the existing hydraulic output settingswhen a pressure droop occurs.
 7. The method of claim 1, wherein saiddrive motor is used both for delivering energy and for reclaiming energyfrom the hydraulic output.
 8. The method of claim 1, wherein thealgorithm includes a subroutine for measuring pump shaft torque outputand translating the measured torque output into a pressure deliveredsignal.
 9. The method of claim 1, wherein the algorithm includes asubroutine for maintaining a constant horsepower from the drive motor,thereby limiting hydraulic output to the application similarly.
 10. Themethod of claim 1, further comprising means for storing electricalenergy including reclaimed energy from regeneration, and for providingfor the requirements of high energy applications typically requiring ahydraulic accumulator.
 11. The method of claim 1, wherein the algorithmincludes a subroutine such that a volumetric pulse correlated with thehydraulic output is used to position the pump cylinders.
 12. The methodof claim 1, wherein the algorithm includes a subroutine for detecting apump leakage rate and outputting an alarm when a predetermined leakagelimit is exceeded.
 13. The method of claim 1, wherein the algorithmincludes a subroutine for assessing a pump output level and applying aprofile of torque vs. velocity corresponding to the assessed outputlevel.
 14. A hydraulic pump system comprising: a hydraulic pump; a drivemotor directly coupled to said pump; and a motor control coupled to saidpump for controlling said drive motor; said motor control employing amotion control algorithm to control a hydraulic output of the pump,wherein said algorithm measures and corrects for any combination of orall three physical parameters of the motor, pump and hydraulic fluid,thereby optimizing an applied torque and resultant pump pressure, andwherein said three physical parameters are windage torque, viscoustorque and coulomb torque and application items REDUX V and REDUX P,respectively.
 15. The system according to claim 14, wherein a pluralityof drives, motors and pumps are electronically line-shafted togetherproviding precise and stable control of pressure and flow being fed to acommon output header regardless of individual pump characteristics orfluid condition variations.
 16. The system according to claim 14,wherein a precise measurement and control of press-ram position, and/orram velocity is achieved.
 17. The system according to claim 14, whereinsaid system monitors and/or controls hydraulic pressures created by thepump, which after mathematical correction for line losses at variousflow rates, mimics load at the press ram.
 18. The system of claim 14,wherein the algorithm includes a subroutine for eliminating a pressurespike; wherein any excess hydraulic output is used to generateelectrical energy when a pressure spike occurs; the electrical energybeing stored in an energy storage means.
 19. The system of claim 14,wherein the algorithm includes a subroutine for eliminating a pumppressure droop; wherein an input signal overrides the existing hydraulicoutput settings when a pressure droop occurs.
 20. The system of claim14, wherein said drive motor is used both for delivering energy and forreclaiming energy from the hydraulic output.
 21. The system of claim 14,wherein the algorithm includes a subroutine for measuring pump shafttorque output and translating the measured torque output into a pressuredelivered signal.
 22. The system of claim 14, wherein the algorithmincludes a subroutine for maintaining a constant horsepower from thedrive motor, thereby limiting hydraulic output to the applicationsimilarly.
 23. The system of claim 14, further comprising means forstoring electrical energy including reclaimed energy from regeneration,and for providing for the requirements of high energy applicationstypically requiring a hydraulic accumulator.
 24. The system of claim 14,wherein the algorithm includes a subroutine such that a volumetric pulsecorrelated with the hydraulic output is used to position the pumpcylinders.
 25. The system of claim 14, wherein the algorithm includes asubroutine for detecting a pump leakage rate and outputting an alarmwhen a predetermined leakage limit is exceeded.
 26. The system of claim14, wherein the algorithm includes a subroutine for assessing a pumpoutput level and applying a profile of torque vs. velocity correspondingto the assessed output level.